1. Field of the Invention
The invention relates to a hydraulic control device and hydraulic control method for a vehicle automatic transmission that engages or releases a hydraulic friction engagement device through a hydraulic pressure applied from an electromagnetic valve device and, more particularly, to a technique for appropriately setting an engagement hydraulic pressure applied to a hydraulic friction engagement device.
2. Description of the Related Art
There is widely known a vehicle automatic transmission in which a plurality of hydraulic friction engagement devices are selectively engaged to establish a plurality of speeds having different speed ratios. For example, Japanese Patent Application Publication No. 10-9377 (JP-A-10-9377) describes such an automatic transmission. JP-A-10-9377 describes that an estimated torque input to the automatic transmission (hereinafter, referred to as transmission input torque) is calculated on the basis of an estimated engine torque calculated and then required engagement hydraulic pressures that are required to keep respective hydraulic friction engagement devices engaged are calculated on the basis of the transmission input torque. In addition, the required engagement hydraulic pressures are, for example, supplied from corresponding linear solenoid valves using a line hydraulic pressure as a source pressure. The linear solenoid valves are provided in correspondence with the hydraulic friction engagement devices. Therefore, the line hydraulic pressure is set to a hydraulic pressure value by which the required engagement hydraulic pressures can be obtained.
FIG. 27 is a view that shows an example of a case where an engagement hydraulic pressure is directly supplied to a hydraulic friction engagement device 2 from a linear solenoid valve 1 without passing through, for example, a shift control valve, or the like. During a non-shift operation in which a predetermined speed of the automatic transmission is kept, the linear solenoid valve 1 is, for example, placed in a regulated state as shown in FIG. 27 in which hydraulic pressure is balanced to thereby supply the required engagement hydraulic pressure to the corresponding hydraulic friction engagement device 2. The regulated state of the linear solenoid valve 1 is, for example, balanced when PC×A+FSP=F is satisfied and is expressed by PC (F−FSP)/A where the engagement hydraulic pressure, that is, the output hydraulic pressure of the linear solenoid valve 1, is PC, the driving force of an electromagnetic valve is F, the reaction force (urging force) of a spring 3 is FSP, and the area of a spool valve element 5, receiving the output hydraulic pressure PC introduced into a feedback oil chamber 4, is A. In consideration of the fact that the output hydraulic pressure PC that exceeds the line hydraulic pressure PL, which is the source pressure, cannot be obtained in terms of the driving characteristic of the linear solenoid valve 1 shown in FIG. 27, when the required engagement hydraulic pressure is set as the line hydraulic pressure PL in a steady state during a non-shift operation, it is only necessary that a hydraulic pressure command value, that is, a driving current, for the linear solenoid valve 1 is set so that the set output hydraulic pressure PC is equal to the line hydraulic pressure PL in order to minimize the power consumption of each linear solenoid valve 1.
Incidentally, in the above steady state in which the line hydraulic pressure PL is set as the output hydraulic pressure PC, for example, when the estimated engine torque varies in accordance with an operating state of the vehicle, the set line hydraulic pressure PL is varied accordingly; however, there is a possibility that the operating state of each linear solenoid valve 1 varies from the above regulated state because of a delay in response of an actual engine torque with respect to the estimated engine torque, variations (individual difference) of the linear solenoid valve 1, or the like. For example, although the hydraulic pressure command value of each linear solenoid valve 1 is supposed to correspond to the line hydraulic pressure PL to form the regulated state shown in FIG. 27, the spool valve element 5 is actually displaced toward the spring 3 with respect to a position in the regulated state to thereby place the linear solenoid valve 1 in a non-regulated state where an input port 6 that introduces the line hydraulic pressure PL is open. Then, during a shift operation of the automatic transmission thereafter, there is a possibility that, when a drain port 7 of the linear solenoid valve 1 is opened to vary the output hydraulic pressure (engagement hydraulic pressure) PC output from the linear solenoid valve 1 toward zero in order to release the hydraulic friction engagement device 2, the response of hydraulic pressure is fast when starting from the regulated state and is slow when starting from the non-regulated state where the input port 6 is open. FIG. 28 is a graph that shows an example of variations in response time (•-• in the graph) and median values of the variations (alternate long and short dashed line in the graph) when the hydraulic friction engagement device 2 is released in a state where the line hydraulic pressure PL is set as the output hydraulic pressure PC in the steady state. In the variations of FIG. 28, the regulated state corresponds to an end of each bar, of which a response time is short, and the non-regulated state where the input port 6 is open corresponds to the other end of each bar, of which a response time is long. In addition, as the source pressure (that is, line hydraulic pressure PL) before the automatic transmission starts shifting decreases, an amplitude from the median value toward the end of each bar, corresponding to the non-regulated state where the input port 6 is open, increases, so variations in response time also increase. In this way, there is a possibility that the operating state of the linear solenoid valve 1 varies to thereby cause a difference in response of the output hydraulic pressure of each linear solenoid valve 1. Therefore, there is a possibility that the response of each engagement hydraulic pressure PC becomes a variable factor to thereby, for example, influence the releasing performance of a release-side hydraulic friction engagement device during a shift operation. In addition, there is a possibility that each linear solenoid valve 1 causes a difference in step response in an initial current at the time of start of shift operation due to an individual difference, or the like, of the linear solenoid valve 1. Thus, robustness of shift operation is lost because of such variations in response. As a result, this may cause an increase in shift shock. Note that it is conceivable that the hydraulic pressure command value of the linear solenoid valve 1 is set so that the set output hydraulic pressure PC becomes the maximum hydraulic pressure of the linear solenoid valve 1 in consideration of variations in terms of design although the output hydraulic pressure PC becomes the line hydraulic pressure PL similarly to thereby reduce the above described variations in response. However, the linear solenoid valve 1 is driven at the maximum power and, therefore, the power consumption of each linear solenoid valve 1 is maximal, so there is a possibility that fuel economy deteriorates. As described above, there has not been suggested that the response of the output hydraulic pressure (engagement hydraulic pressure) is stabilized while saving the power consumption of each linear solenoid valve.